JP2005532523A - Method and apparatus for detecting flash gas - Google Patents

Abstract

A method and a device for detecting flash gas in a vapor-compression refrigeration or heat pump system comprising a compressor, a condenser, an expansion device, and an evaporator interconnected by conduits providing a flow path for a refrigerant, by determining a first rate of heat flow of a heat exchange fluid flow across a heat exchanger of the system and a second rate of heat flow of the refrigerant across the heat exchanger, and using the rates of heat flow for establishing an energy balance from which a parameter for monitoring the refrigerant flow is derived, to thereby provide early detection of flash gas with a minimum number of false alarms.

Description

  The present invention relates to a method for detecting flash gas in a vapor compression refrigeration system or heat pump system including a compressor, a condensing device, an expansion device, an evaporator, interconnected by conduits providing a refrigerant flow path. The present invention relates to a flash gas detection device.

  In a vapor compression refrigeration system or heat pump system, refrigerant circulates through the system and undergoes phase changes and pressure changes. In this system, the refrigerant gas is compressed in a compressor in order to obtain a high-pressure refrigerant gas, which is supplied to a condenser (heat exchanger) where the refrigerant gas is cooled and liquefied. When exiting the condenser, the refrigerant is in a liquid state and expands to a low pressure in the expansion device and evaporates this refrigerant in the evaporator (heat exchanger) to obtain a low pressure refrigerant gas The processing step continues as it is fed to the compressor.

  However, in some cases, gas phase refrigerant is present in the liquid refrigerant conduit due to boiling liquid refrigerant. This refrigerant gas in the liquid refrigerant conduit is labeled “flash gas”. The presence of flash gas at the inlet to the inflator greatly reduces the capacity of the inflator flow due to the clogging effect of the inflator, thus compromising system efficiency. This effect can result in the system using more energy than necessary and possibly not heating or cooling as expected, for example a store display cabinet could lead to warming the food in the cabinet, and hence The food will have to be dumped. In addition, the system components are outside the normal operating exterior. Due to the high load and low refrigerant flow, the compressor can be exposed to overheating when flash gas is present, especially the mist oil in the refrigerant is expected to function as a lubricant In some cases, the compressor will run out of lubricant, causing the compressor to seize.

  Flash gas is caused by several factors. 1) The condenser cannot liquefy all the refrigerant due to the high temperature of the heat exchanger fluid. 2) The coolant level is low due to insufficient charge or leakage. 3) The system is not properly designed, for example there is a relatively long conduit without insulation from the condenser to the expansion device, the refrigerant re-heats and evaporates, or there is a relatively large pressure drop in the conduit Present and the refrigerant evaporates.

  Leakage in the system is a serious problem because the refrigerant selected is detrimental to human or animal health or the environment. In particular, some refrigerants are suspected to contribute to the ozone depletion process. In any case, refrigerants are very expensive and are often heavily taxed, so the system recharge for a normal refrigerated display cabinet for a store is quite expensive. Recently, at a store with a refrigerated display cabinet, half of the refrigerant was lost before the refrigeration system was found to be leaking, but the system recharge was 75,000 dkr, at a cost of about $ 10,000.

  A known method for detecting flash gas is to provide a viewing window in the liquid conduit of the system so that bubbles in the liquid can be observed. This wastes labor and time, and the observation of bubbles can lead to the wrong direction, as even small amounts of bubbles can sometimes be present even in a well-functioning system.

  Another method is to indirectly detect the flash gas by triggering an alarm when the expansion device is fully open, such as when the expansion device is an electronic expansion valve. In this case, a significant number of false alarms can be experienced as it can also occur in a system where a well-developed expansion device is functioning properly without flash gas.

  It is an object of the present invention to provide a method for early detection of flash gas with a minimal number of malfunction alarms.

  The purpose is to determine the first rate of heat flow of the heat exchange fluid flow across the heat exchanger of the system and the second rate of heat flow of the refrigerant across the heat exchanger and derive parameters for monitoring the refrigerant flow. Filled by a method that includes using those heat flow rates to achieve the desired energy balance. This allows the refrigerant flow to be monitored without direct measurement using a flow meter. Such flow meters are expensive and can further limit the flow rate.

  According to one embodiment, the heat exchanger is an evaporator, which is an ideal component.

  According to an alternative or additional embodiment, the heat exchanger is a condenser.

  As one skilled in the art will appreciate, the first rate of heat flow of the heat exchange fluid can be obtained in a variety of ways, but according to one embodiment, the method involves the flow of heat exchange fluid and the heat across the heat exchanger. The first rate of heat flow is obtained by obtaining the specific enthalpy change of the exchange fluid.

  According to one embodiment, the method includes obtaining the flow rate of the heat exchange fluid as a constant based on empirical data or data obtained from fault-free operation of the system.

  According to one embodiment, the method includes obtaining a specific enthalpy change of the heat exchange fluid across the heat exchanger based on the temperature measurement of the heat exchange fluid before and after the heat exchanger.

  The second rate of the refrigerant heat flow can be determined by obtaining the refrigerant flow rate and the change in the specific enthalpy of the refrigerant across the heat exchanger.

  The flow rate of the refrigerant can be obtained in a variety of ways, including direct measurement, but this is not preferred. According to one embodiment, the method comprises the flow characteristics of the expansion device, the expansion device open flow path and / or the open period, the absolute pressure before and after the expansion device, and if necessary the refrigerant flow at the expansion device inlet. Obtaining the flow rate of the refrigerant based on some supercooling.

  The difference in specific enthalpy of the refrigerant flow is obtained based on the record of the refrigerant temperature and pressure at the inlet of the expansion device, or the record of the refrigerant evaporator outlet temperature and the refrigerant evaporator outlet pressure or the refrigerant saturation temperature at the evaporator inlet. It is possible.

  A direct assessment of the refrigerant flow rate is possible, but it can suffer from some disadvantages, for example, due to parameter variations or fluctuations within the refrigeration system or heat pump system, and thus the method is Preferably, obtaining a residual as a difference between the first rate and the second rate of heat flow.

  In order to reduce the sensitivity to parameter variations and fluctuations in the system, and to be able to record the refrigerant flow trend initially, the method includes providing a failure indicator by the remainder. formula

によって与えられ、ここでＳ_μ1，iは次式

Where S _{μ1, i} is given by

に従って計算され、ここで
ｒ_i：残余
ｋ₁：比例定数
μ₀：第１の感度値
μ₁：第２の感度値
である。

Where r _i : residual k ₁ : proportional constant μ ₀ : first sensitivity value μ ₁ : second sensitivity value.

  According to a second aspect, the present invention relates to a flash gas detection device, comprising: a first rate of heat flow of a heat exchange fluid stream across a heat exchanger of the system; and a second rate of heat flow of a refrigerant across the heat exchanger. Means for determining, and means for using the rate of heat flow to obtain an energy balance from which parameters for monitoring the refrigerant flow are derived, the apparatus further evaluating the refrigerant flow rate to produce an output signal Evaluation means for including.

  According to one embodiment of the apparatus, the means for determining the first rate of heat flow includes means for sensing the temperature of the heat exchange fluid before and after the heat exchanger.

  According to one embodiment of the apparatus, the means for determining the second rate of heat flow is means for sensing the temperature and pressure of the refrigerant at the inlet of the expansion device, and for sensing the temperature of the refrigerant at the outlet of the evaporator. Means for obtaining the pressure or saturation temperature at the outlet of the expansion device.

  According to one embodiment of this device, the means for obtaining the second rate of heat flow obtains means for sensing the absolute pressure of the refrigerant before and after the expansion device, the expansion device open flow path or the open period. Means for.

  In order to provide a robust evaluation means, the evaluation means comprises a first value comprised of the flow rate of the heat exchange fluid flow and the specific enthalpy change across the system heat exchanger, the flow rate of the refrigerant and the heat of the system. Means for obtaining a remainder as a difference between a second value constituted by the specific enthalpy change of the refrigerant across the exchanger.

  In order to be able to evaluate trends in the output signal, the apparatus further comprises memory means for storing the output signal, means for comparing the output signal with the previously stored output signal. Is possible.

  In the following, the present invention will be described in detail with reference to the drawings.

  Although discussed below for a simple refrigeration system, the principles apply equally to heat pump systems and the present invention is in no way limited to refrigeration systems, as will be appreciated by those skilled in the art.

  A simple refrigeration system is shown in FIG. The system includes a compressor 5, a condenser 6, an expansion device 7, and an evaporator 8 interconnected by a conduit 9 through which refrigerant flows. The mode of operation of the system is well known and the compression and condenser 6 from the temperature and pressure at point 1 in front of the compressor 5 to the higher temperature and pressure at point 2 behind the compressor 5. Includes condensation of the refrigerant by heat exchange with the heat exchange fluid in the condenser 6 to obtain a liquid refrigerant at high pressure at point 3 behind. The high-pressure refrigerant liquid expands in the expansion device 7 and becomes a mixture of low-pressure liquid and gaseous refrigerant at point 4 behind the expansion device. In this simple example, the expansion device is an expansion valve, but other types of expansion devices such as turbines, orifices, or capillary tubes are possible. Behind the expansion device, the mixture flows into the evaporator 8 where the liquid is vaporized by heat exchange with the heat exchange fluid in the evaporator 8. In this simple example, the heat exchange fluid is air, but this principle applies equally to refrigeration systems or heat pump systems that use other heat exchange fluids, such as salt water, as well as the heat in the condenser and evaporator. The exchange fluid need not be the same.

  FIG. 2 is a diagram of log p, h of the refrigeration system according to FIG. 1, showing the refrigerant pressure and enthalpy. Reference numeral 10 represents a saturated vapor curve, 11 represents a saturated liquid curve, and C.I. P. Is the critical point. In the region 12 to the right of the saturated vapor line 10, the refrigerant is a superheated gas, while in the region 13 to the left of the saturated liquid line 11, the refrigerant is a supercooled liquid. In region 14, the refrigerant is a mixture of gas and liquid. As can be seen, at point 1 in front of the compressor, the refrigerant is completely gaseous, and during compression the refrigerant pressure and temperature are raised, so at point 2 behind the compressor, the refrigerant is at high pressure. Is a superheated gas. The refrigerant leaving the condenser 6 at point 3 should be completely liquid, i.e. the refrigerant should be in a state above the saturated liquid curve 11 or in the region 13 of supercooled liquid refrigerant. The refrigerant is expanded in the expansion device 7 and becomes a mixture of low-pressure liquid and gas at a point 4 behind the expansion device 7. In the evaporator 8, the refrigerant is vaporized at a constant pressure by heat exchange with the heat exchange fluid, thereby becoming completely gaseous at the point 1 at the outlet of the evaporator.

  If the refrigerant entering the expansion device 7 is a liquid and a gas, that is, a mixture of the aforementioned flash gas, as indicated by the point 3 ', the flow rate of the refrigerant is limited as described above and the cooling capacity of the evaporator of the refrigeration system Is greatly reduced. Furthermore, the available enthalpy difference in the evaporator 8 is not so small, which again reduces the cooling capacity.

  FIG. 3 schematically shows a refrigerated display cabinet having a refrigeration system. Refrigerated display cabinets are used in supermarkets to display and sell refrigerated or frozen foods. The refrigerated display cabinet has a storage compartment 15 in which food is stored. An air passage 16 is arranged around the storage compartment 15, ie the air passage 16 continues on both sides and below the storage compartment 15. The air flow 17 indicated by the arrow passes through the air passage 16 and then enters the cooling zone 18 above the cooling section 15. The air is then again led to the inlet to the air passage 16. There is a mixing zone. Within the mixing zone 19, the air stream 17 is mixed with ambient air. This replenishes the air that has entered the storage compartment or has somehow escaped. A blower device 20 composed of one or a plurality of fans is provided in the air passage 16. The air flow 17 can be reliably moved in the air passage 16 by the blower 20. Since the system evaporator 8 is located in the air passage 16, the refrigerated display cabinet includes the components of a simple refrigeration system as shown schematically in FIG. The evaporator 8 is a heat exchanger that exchanges heat between the refrigerant in the refrigeration system and the air flow 17. In the evaporator 8, the refrigerant draws heat from the air stream 17, so that the air stream is cooled. The cycle of this refrigeration system is as described in FIGS. 1, 2 and the numbers used for them.

  As explained, it would be very advantageous in a refrigeration system or heat pump system if the presence of flash gas or gas at the inlet of the expander could be detected. The effect of the flash gas is a decrease in the flow rate through the expansion device at the expansion device inlet when compared to the normal situation flow rate of liquid-only refrigerant. This difference is an indication of the presence of flash gas when the refrigerant flow in the refrigeration system is less at the expansion device inlet than the theoretical refrigerant flow provided by the liquid phase refrigerant. It is possible to obtain the refrigerant flow rate by direct measurement using a flow meter. However, such flow meters are relatively expensive, and further restricting the flow rate can cause a pressure drop, which in itself can lead to the formation of flash gas, ensuring system efficiency. To lose. Therefore, it is preferable to obtain the refrigerant flow rate by other means. One possible way is to obtain the refrigerant flow rate based on the principle of energy conservation or energy balance of one of the heat exchangers of the refrigeration system, namely the evaporator 8 or the condenser 6. The following will be discussed with respect to the evaporator 8, but the person skilled in the art will understand that the condenser 6 can be used as well.

  The energy balance of the evaporator 8 is based on the following equation.

ここで

here

は単位時間あたりに空気から除去される熱、すなわち空気から引き渡される熱流のレートであり、

Is the heat removed from the air per unit time, ie the rate of heat flow delivered from the air,

は単位時間あたりに冷媒によって吸い取られる熱、すなわち冷媒へ引き渡される熱流のレートである。

Is the heat absorbed by the refrigerant per unit time, that is, the rate of heat flow delivered to the refrigerant.

  Refrigerant heat flow rate

すなわち単位時間あたりに冷媒へと渡される得るための基礎は次の式

That is, the basis for getting passed to the refrigerant per unit time is

であり、ここで

And here

は冷媒の流量である。ｈ_Ref，Outは蒸発器出口での冷媒の比エンタルピであり、ｈ_Ref，Inは蒸発器入口での冷媒の比エンタルピである。冷媒の比エンタルピは冷媒の材料と状態の特性であり、比エンタルピはどのような冷媒についても決定することが可能である。冷媒の製造業者は冷媒に関する図２によるタイプのｌｏｇ ｐ、ｈ図を提供する。この図によって、蒸発器を横切る比エンタルピの差を得ることが可能である。例えばｌｏｇ ｐ、ｈ図を用いてｈ_Ref，Inを得るためには、膨張装置入口での冷媒の温度と圧力（それぞれＴ_Ref，InとＰ_Con）を知ることが必要とされるだけである。それらのパラメータを、温度センサまたは圧力センサによって測定することができる。蒸発器８と冷蔵システムの測定ポイント、パラメータ測定ポイント、パラメータを、図３による冷蔵陳列キャビネットの一部を示す略図である図４に見ることが可能である。

Is the flow rate of the refrigerant. h _{Ref, Out} is the specific enthalpy of the refrigerant at the evaporator outlet, and h _{Ref, In} is the specific enthalpy of the refrigerant at the evaporator inlet. The specific enthalpy of the refrigerant is a property of the material and state of the refrigerant, and the specific enthalpy can be determined for any refrigerant. The manufacturer of the refrigerant provides a log p, h diagram of the type according to FIG. 2 for the refrigerant. This figure makes it possible to obtain the specific enthalpy difference across the evaporator. For example _, to obtain h _{Ref and In} using log p and h diagrams, it is only necessary to know the temperature and pressure of the refrigerant at the inlet of the expansion device (T _{Ref, In} and P _{Con ,} respectively). . These parameters can be measured by temperature sensors or pressure sensors. The measurement points, parameter measurement points and parameters of the evaporator 8 and the refrigeration system can be seen in FIG. 4, which is a schematic diagram showing a part of the refrigerated display cabinet according to FIG.

Two measurements are required to obtain the specific enthalpy at the evaporator outlet. That is, the temperature at the outlet of the evaporator (T _{Ref, out} ) and the pressure at the outlet (P _{Ref, out} ) or the saturation temperature (T _{Ref, sat} ). The temperature at the outlet of the evaporator 8 can be measured with a temperature sensor, and the pressure at the outlet can be measured with a pressure sensor.

  It is of course possible to use values obtained from charts or tables instead of log p, h diagrams, which simplify the calculation with the help of the processor. Refrigerant manufacturers often also provide state equations for the refrigerant.

The flow rate of the refrigerant is obtained by assuming a liquid-phase refrigerant at the inlet of the expansion device. For example, in a refrigeration system with an electronically controlled expansion valve using pulse width modulation, if the absolute pressure difference across the valve and the subcooling (T _{V, in} ) at the expansion valve inlet are known, the valve open flow It is possible to determine the theoretical refrigerant flow based on the path and / or the open period. Similarly, refrigerant flow rates can be obtained in refrigeration systems that use expansion devices with known open flow paths, such as fixed orifices or capillary tubes. In most systems, the above parameters are already known because there is a pressure sensor that measures the pressure in the condenser 6. In many cases, the supercooling is almost constant and slight and can be estimated, so there is no need to measure the supercooling. It is then possible to calculate the theoretical refrigerant flow through the expansion valve according to valve characteristics, pressure differential, degree of supercooling, valve open flow path and / or valve open period. In many pulse width modulation type expansion valves, it is known for constant supercooling that the theoretical refrigerant flow rate is approximately proportional to the difference between the absolute pressures before and after and the open period of the valve. In this case, the theoretical flow rate is calculated according to the following formula:

ここでＰ_Conは凝縮器内の絶対圧力であり、Ｐ_Ref，outは蒸発器内の圧力であり、ＯＰは開放期間であり、ｋ_Expは弁と過冷却に応じて決まる比例定数である。いくつかのケースでは、膨張弁を通って流れる冷媒が過冷却によって影響を受けるほど冷媒の過冷却度が大きいので、過冷却度を測定する必要がある。しかしながら多くのケースでは、過冷却度は小さく、かつ殆ど一定であるので、弁の前後の絶対圧力と弁の開放流路および／または開放期間を得ることだけが必要とされ、過冷却度は弁の特性または比例定数の中で考慮することが可能である。

Here, P _Con is the absolute pressure in the condenser, P _{Ref, out} is the pressure in the evaporator, OP is the open period, and k _Exp is a proportional constant determined according to the valve and _subcooling . In some cases, the degree of supercooling needs to be measured because the degree of supercooling of the refrigerant is so great that the refrigerant flowing through the expansion valve is affected by the supercooling. However, in many cases, the degree of supercooling is small and almost constant, so it is only necessary to obtain the absolute pressure before and after the valve and the open flow path and / or duration of the valve. Can be considered in the characteristics or proportionality constants.

  Similarly, air heat flow rate

すなわち単位時間あたりに空気によって吸い取られる熱は式

That is, the heat absorbed by the air per unit time is the formula

に従って得られ、ここで

Obtained here and here

は単位時間あたりの空気の流量であり、ｈ_Air，inは蒸発器の前での空気の比エンタルピであり、ｈ_Air，outは蒸発器の後ろでの空気の比エンタルピである。

Is the flow rate of air per unit time, h _{Air, in} is the specific enthalpy of air before the evaporator, and h _{Air, out} is the specific enthalpy of air behind the evaporator.

The specific enthalpy of air can be calculated based on the following formula:
h _Air = 1.006 x t + x (2501 + 1.8 x t), [h] = kJ / kg (5)
Here, t is the temperature of the air, that is, T _{Eva, in} before the evaporator, and T _{Eva, out} after the evaporator. x represents the absolute humidity of the air. The absolute humidity of air is

によって計算できる。

Can be calculated by

Here, P _W is the partial pressure of water vapor in the air, and P _Amb is the pressure of air. Either P _Amb is measured, or standard atmospheric pressure is simply used. The actual pressure difference obtained from standard atmospheric pressure is not very important for the calculation of the amount of heat per unit time delivered by air. The partial pressure of the water vapor is determined by the relative humidity of the air and the saturated water vapor pressure, and the following formula: P _W = P _{W, Sat} × RH (7)
Can be calculated by

Here, RH is the relative humidity of air, and P _{W, Sat} is the saturation pressure of water vapor. P _{W and Sat} alone depend on temperature and can be found from thermodynamic reference books. The relative humidity of the air can be measured, or the normal value can be used for the calculation.

  When the expressions (2) and (4) are equal as shown in the expression (1), the following expression is obtained.

  Air flow rate from now on

は

Is

を

The

と分離することによって得ることができる。

And can be obtained by separating them.

  Given the correct air flow, this equation can be used to evaluate system operation.

  In many cases, it is recommended to record the theoretical flow rate of air in the system. For example, this theoretical air flow rate can be recorded as an average over a specific time period during which the refrigeration system is operating in stable and normal operating conditions. Such a time period may be, for example, 100 minutes.

  One challenge lies in the fact that signals coming from various sensors (thermometers, pressure sensors) are subject to significant fluctuations. These variations can result in signals regarding the theoretical refrigerant flow as the opposite phase, which poses some challenges for analysis. These variations or variations are the result of kinetic conditions within the refrigeration system. Therefore, periodically, for example once a minute, the value indicated as “residual” in the following equation based on the energy balance according to equation (1)

を得ることが有利である。式（２）と（４）に基づくと残余は

It is advantageous to obtain Based on equations (2) and (4), the remainder is

として得ることが可能であり、ここで

As can be obtained here

は見積もられた空気の流量であって上述のように、すなわち正常な動作の期間の平均値として得られる。他の可能性は

Is the estimated air flow and is obtained as described above, ie, as an average value during normal operation. Other possibilities are

が一定の値であり、一定に作動する送風機を有する図３、４のような冷蔵陳列キャビネットの極めて単純な例で得られることが可能である。

Is a constant value and can be obtained in a very simple example of a refrigerated display cabinet as shown in FIGS.

  In a refrigeration system that operates without problems, the residual r varies considerably, but on average it is zero. In order to enable early detection of faults indicated as trends in the residual, whether the value recorded for the residual r follows a Gaussian distribution with an average value, whether the refrigeration system is working without fault or whether a fault has occurred Is assumed to be irrelevant.

  Since the course's principle of energy conservation or energy balance is invariant, the remainder should be zero in principle, regardless of whether there is a fault in the system. If this is not the case in the above formula, it is because the preconditions for the use of the formula used are not met in the case of a system malfunction.

When flash gas is present in the expansion device, the valve characteristics change so that k _Exp is several times smaller. This is not taken into account in the calculation and hence the rate of refrigerant heat flow used in the formula

は現実よりもはるかに大きくなる。空気の熱流のレート

Is much larger than reality. Air heat flow rate

に対しては、（熱交換器を横切る空気流の減少を引き起こす不具合が生じなかったと仮定すると）計算は正しく、それは熱交換器を横切る空気の熱流のレート

The calculation is correct (assuming no failure has occurred that would cause a reduction in airflow across the heat exchanger), which is the rate of air heat flow across the heat exchanger.

について算出された値が現実の空気の熱流のレートに等しいことを意味する。結論は、膨張装置内にフラッシュガスがある場合には残余の平均が負になるということである。

Means that the value calculated for is equal to the actual air heat flow rate. The conclusion is that if there is flash gas in the expansion device, the residual average will be negative.

  If there is a defect (defective fan or freezing of the heat exchanger) that causes a reduction in the air flow across the heat exchanger, the air flow is a value related to the air flow used in the calculation.

よりも小さくなる。これは、計算に使用される空気の熱流のレートが現実の空気の熱流の実際のレートよりも大きいこと、すなわち空気から除去される単位時間あたりの熱が予期される値よりも少ないことを意味する。結論は、（冷媒の熱流のレートが正しい、すなわちフラッシュガスがないとを仮定すると）熱交換器を横切る空気流の減少を引き起こす不具合があれば残余が正になるということである。

Smaller than. This means that the rate of air heat flow used in the calculation is greater than the actual rate of real air heat flow, i.e., less heat per unit time is removed from the air than expected. To do. The conclusion is that the remainder will be positive if there is a fault that causes a reduction in the air flow across the heat exchanger (assuming that the rate of heat flow of the refrigerant is correct, ie no flash gas).

To filter out the residual signal for any fluctuations or vibrations, a statistical operation is performed by examining the following hypothesis.
1. The average value of the residual r is μ ₁ (μ ₁ <0). Compatible with flash gas testing.
2. The average value of the residual r is μ ₂ (μ ₂ > 0). Supports testing for reduced airflow.

This investigation is done by calculating two failure indicators according to the following formula:
1. Test against flash gas

ここでＳ_μ1，iは次の式によって計算される。

Here, S _{μ1, i} is calculated by the following equation.

ここでｋ₁は比例定数であり、μ₀は第１の感度値であり、μ₁は第２の感度値であって上記で示されるように負であり、
２．低下した空気流に対するテスト

Where k ₁ is a proportionality constant, μ ₀ is the first sensitivity value, μ ₁ is the second sensitivity value and is negative as indicated above,
2. Test for reduced airflow

ここでＳ_μ2，iは次の式によって計算される。

Here, S _{μ2, i} is calculated by the following equation.

ここでｋ₁は比例定数であり、μ₀は第１の感度値であり、μ₂は第２の感度値であって上記で示されるように正である。

Where k ₁ is a proportionality constant, μ ₀ is the first sensitivity value, μ ₂ is the second sensitivity value, and is positive as indicated above.

In the equation (11), that is, at the first point in time, it is naturally assumed that the failure index S _{μ1, i} is set to zero. For the later point in time, S _{μ1, i} is used according to the equation (12), and the sum of this value and the failure index S _{μ1, i} at the previous point in time is calculated. When this sum is greater than zero, the failure indicator is set to this new value. When this sum is less than or equal to zero, the failure index is set to zero. In the simplest case, μ ₀ is set to zero. μ ₁ is a selected value obtained, for example, when a malfunction occurs. The parameter μ ₁ is a decision criterion regarding the frequency with which it is permissible to receive fault alarms regarding the detection of flash gas.

Similarly, in the equation (13), it is naturally assumed that the failure index S _{μ2, i} , that is, the first point in time is set to zero. For the later point in time, S _{μ2, i} is used according to the equation (14), and the sum of this value and the failure index S _{μ2, i} at the previous point in time is calculated. When this sum is greater than zero, the failure indicator is set to this new value. When this sum is less than or equal to zero, the failure index is set to zero. In the simplest case, μ ₀ is set to zero. μ ₂ is an estimated value obtained when, for example, a failure occurs. The parameter μ ₂ is a determination criterion regarding the frequency with which it is permitted to receive a malfunction alarm regarding the air flow rate.

For example, if a failure occurs where flash gas is present at the inlet of the expansion valve, the failure index will increase because the average S _{μ1, i} value recorded regularly is greater than zero. When the failure index reaches a predetermined value, an alarm is activated, and the alarm indicates that the flow rate of the refrigerant is reduced. If smaller value, i.e. if more negative mu ₁ is selected, it becomes even less malfunction alarm, but the sensitivity is a risk of falls exists regarding defect detection.

The principle of the filtering operation according to equations (11) and (13) is specifically illustrated by FIGS. In FIG. 5, the time in minutes is on the x-axis and the remainder r is on the y-axis. There was a blower fault between t = 200 and 300 minutes, which caused a significant increase in the remainder. Further, there is a flash gas between t = 1090 to 1147 and 1455 to 1780, which is seen as a significant reduction in the residual to a value of about −10 × 10 ⁶ . However, as can be seen, the signal is not exposed to significant variations and fluctuations that make evaluation difficult.

FIG. 5 shows a situation of a different defect. If the defect indicators S _{μ1, i} and S _{μ2, i} whose behavior is seen in FIG. 6 are monitored, there is a possibility of better identification. Here, a broken line indicates S _{μ1, i,} and a continuous line indicates S _{μ2, i} . Here, the values of the defect indices S _{μ1, i} , S _{μ2, i} are on the y axis, and the time in minutes is on the x axis. Due to the failure of the blower, the failure index S _{μ2, i} continuously increases between t = 200 and 330 minutes. An alarm will be triggered if S _{μ2, i} exceeds a value of 0.2 × 10 ⁹ , for example. As can be seen from the comparison between FIGS. 5 and 6, early detection is possible particularly when this defect index is used. Similarly, when the failure index S _{μ1, i} rises between t = 1090 and 1147 due to the flash gas, then gradually decreases to zero, and then the flash gas exists again at the expansion valve inlet. It rises again in the period from t = 1455 to 1780. If the refrigeration system continues to operate satisfactorily for a long period of time, the failure indices S _{μ1, i} , S _{μ2, i} are _reset to zero. In practice, if the defect is corrected, the defect index S _{μ1, i} , S _{μ2, i} will be set back to zero.

As can be seen in FIGS. 5 and 6, it is possible to evaluate the system simultaneously for the reduced air flow and the flash gas at the expander inlet by evaluating the failure index using decision criteria μ ₁ and μ _2. .

  Furthermore, the method and apparatus according to the present invention can provide valuable information regarding the design of refrigeration systems. Many refrigeration systems are tailored specifically for specific applications, such as stores with one or more refrigerated display cabinets, and these refrigeration systems may not be optimal. That is, the conduit is long and the pressure is reduced due to bending of the conduit, or the conduit is exposed to ambient heat. With the present method and apparatus, it is possible to detect that a refrigeration system is not optimal, to evaluate the system and suggest system improvements and / or to propose improvements for future systems It is possible to dispatch specialists.

  A further advantage of the device is that any refrigeration system or heat pump system can be modified without any extensive intervention in the refrigeration system. The device usually uses signals coming from sensors already in the refrigeration system, or sensors that can be retrofitted at a very low cost.

  In the foregoing description, a simple example has been used to illustrate the principles of the present invention, but as will be readily appreciated by those skilled in the art, the present invention is not limited to a plurality of heat exchangers, ie, a plurality of condensers. It can be applied to a complex system having an evaporator and / or multiple evaporators.

1 is a schematic diagram illustrating a simple refrigeration system or heat pump system. Fig. 2 is a schematic log p, h diagram for the cycle of the system according to Fig. 1; 2 is a schematic diagram illustrating a refrigerated display cabinet having the refrigeration system according to FIG. 1. Fig. 4 is a schematic view showing a part of the refrigerated display cabinet according to Fig. 3. It is a figure which shows the remainder of a malfunction state. It is a figure which shows the parameter | index of the malfunction of the malfunction state by FIG.

Claims (17)

  A method for detecting flash gas in a vapor compression refrigeration system or a heat pump system, including a compressor, a condenser, an expansion device, an evaporator, interconnected by conduits forming a refrigerant flow path, the method An energy derived parameter for determining a first rate of heat flow of the heat exchange fluid flow across the heat exchanger of the system and a second rate of heat flow of the refrigerant across the heat exchanger and monitoring the refrigerant flow; Using the rate of heat flow to achieve a balance.   The method of claim 1, wherein a heat exchanger is the evaporator.   The method of claim 1, wherein a heat exchanger is the condenser.   4. A method according to any one of the preceding claims, characterized in that the first rate of heat flow is obtained by obtaining a change in the heat exchange fluid flow rate and the specific enthalpy of the heat exchange fluid across the heat exchanger. .   5. The method of claim 4, comprising obtaining the heat exchange fluid flow rate as a constant based on empirical data or data obtained from faultless operation of the system.   6. A method according to claim 4 or 5, characterized in that the specific enthalpy change of the heat exchange fluid across the heat exchanger is obtained based on the measurement of the temperature of the heat exchange fluid before and after the heat exchanger.   7. The method according to claim 1, wherein the second rate of the refrigerant heat flow is obtained by obtaining the refrigerant flow rate and the specific enthalpy change of the refrigerant across the heat exchanger.   Based on the flow characteristics of the expansion device, the expansion device open flow path and / or the open period, the absolute pressure before and after the expansion device, and if necessary, any refrigerant subcooling at the expansion device inlet The method according to claim 7, wherein a flow rate is obtained.   Obtaining the difference in specific enthalpy of refrigerant flow based on the refrigerant temperature and pressure records at the expansion device inlet and the refrigerant evaporator outlet temperature and refrigerant evaporator outlet pressure or refrigerant saturation temperature at the evaporator inlet. A method according to claim 7 or 8, characterized in that   10. A method according to any one of the preceding claims, characterized in that the residue is obtained as the difference between the first rate of heat flow and the second rate of heat flow. The remainder gives an index of failure,

Where S _{μ1, i} is given by

Is calculated according to where r _i: Residual k _1: proportionality constant mu _0: a first sensitivity value mu _1: The method of claim 10 which is a second sensitivity value.   A flash gas detection device for a vapor compression refrigeration system or heat pump system, including a compressor, a condenser, an expansion device, an evaporator, interconnected by conduits forming a refrigerant flow path, wherein the heat exchange of said system Means for determining a first rate of heat flow of the heat exchange fluid stream across the heat exchanger and a second rate of heat flow of the refrigerant across the heat exchanger, and energy derived parameters for monitoring the refrigerant flow An apparatus comprising: means for using said rate of heat flow to achieve a balance; and evaluating means for evaluating the flow rate of the refrigerant to produce an output signal.   The apparatus of claim 12, wherein the means for determining a first rate of heat flow includes means for sensing the temperature of the heat exchange fluid before and after the heat exchanger.   The means for determining the second rate of heat flow includes means for sensing refrigerant temperature and pressure at the expander inlet and means for obtaining pressure or saturation temperature at the expander outlet. The apparatus according to claim 12 or 13.   Means for determining a second rate of heat flow includes means for sensing absolute refrigerant pressure before and after the expansion device, and means for obtaining an expansion flow path or period of expansion device. 15. A device according to any one of claims 12 to 14, characterized in that it is characterized in that   A first value comprised of the flow rate of the heat exchange fluid flow and the specific enthalpy change across the heat exchanger of the system; and the change means specific enthalpy of the refrigerant across the heat exchanger of the system. 16. Apparatus according to any one of claims 12 to 15, comprising means for obtaining a residual as a difference between a second value constituted by a quantity.   17. Apparatus according to any one of claims 12 to 16, further comprising memory means for storing an output signal and means for comparing the output signal with a previously stored output signal.

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